Wednesday, December 19, 2012

The Rainforest

Recently, I was asked to assist in the design for an enclosed rainforest exhibit. The project has a lot of glass and is in a very southern climate zone (near the Gulf of Mexico). I have looked at a couple of other similar projects in the past and, of course, volunteered to assist. As a preliminary, we managed to get a tour of the Dallas Aquarium, which includes a large enclosed rainforest. 

We were met by the head of maintenance, who seemed to have worked there for 20 years and knew all the history of the HVAC system. Much of the air distribution system was custom built by the facilities folks themselves, in an effort to make the air outlets nearly invisible. I made several general observations concerning projects of this size.

Vents at the top of the structure are essential. At the Dallas facility, we were told that while the vents at the top of the tall glass wall were able to be opened and closed, they just leave them open year round. This seems to be a pretty general recommendation, as all tall structures exhibit stack effect, and releasing the hot air always seems to assist in controlling the space temperatures.

Clear glass seems unwise, especially on the roof. Having the east wall transparent seems to provide sufficient UV for the plants to thrive. It is unclear how much UV passes through the translucent panels used throughout most structures of this type. In general, it seems that displacement ventilation is a natural candidate for these types of spaces, especially as the use of relatively warm water (62°F) to the fan coils results in 65-68°F air at the outlets.

It was very interesting to see how the facilities folks managed the systems and how they often created their own air distribution devices.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Monday, December 10, 2012

Free Jets

I’ve had a couple of inquiries about the throw of air jets, where engineers have asked why some manufacturer’s data is different than others. In truth, a jet of air is pretty much independent of the shape of the hole from which it is discharged. Long narrow jets behave differently than round ones, of course, but there is little actual difference between round jets from different manufacturer’s at a given flow rate. The old Air Diffusion Council (ADC) published a “typical” grille throw chart; most manufacturers use this chart, as it has proven to be close to their measured data. However, there is a huge difference with the performance of the jet with regard to “free” or “entrained” jet performance.

A basic fact of air jet dynamics is that a jet of air has negative static pressure. This results from the laws of conservation of energy and the translation of air from being restrained in a duct and entering a non-constrained space. As the sum of velocity and static pressure must be essentially constant, the total pressure of air in a duct, when discharged into an unrestrained space, results in most of the potential fan energy being translated into velocity pressure. The only way for the sum be constant is for the static pressure to be negative. This is why a jet of air induces air from the surrounding environment. The result is that the mass of moving air increases while the velocity decreases to maintain a conservation of energy. 

A jet that is discharged parallel to a surface experiences two phenomenon. The higher pressure of the air opposite the surface “pushes” the jet towards the surface. The rate of induction is reduced by the reduction in the exposed jet surface, where it is not touching the surface. In practice, an “entrained” jet has about 30% less exposed surface than a “free” jet. If one looks closely at the math, the ratio of most free jets to entrained jets is about the square root of 2, or 0.707.

Krueger, along with most other manufacturers, displays throw for most products as “entrained” jet throws. The ADC “standard” grille throw graph was, in fact, determined along the floor of a large warehouse many years ago. There are a few exceptions however. Drum louvers are seldom installed in a manner where the throw would be along a surface, so we present “free” jet throw data. The same is true of our round jet outlets and some vertical throw data for adjustable outlets. For duct mounted grilles, we provide both free and entrained jet throw data. Be sure to look at the data notes at the bottom of any manufacturer’s throw data to be sure what is being reported.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Thursday, December 6, 2012

ASHRAE Journal Article

In case you missed it, I have part one of a two-part article published in the November ASHRAE Journal. Part II will be in the December issue. I have been giving a talk, from which the articles are based, at ASHRAE chapters and engineer’s offices around the country for the past couple of years. The article’s premise, titled “Methods of Effective Room Air Distribution”, is that folks are responding to the request for more energy efficient systems by designing air distribution systems around alternates to overhead air delivery, which has been the design of choice for the past 30 years. Unlike overhead systems which have many years of data validation, the energy calculations for these newer systems have not been very well vetted. The result is that engineers must modify the inputs or calculations to come up with energy use values. The output, of course, is merely a guess. Hopefully it is an educated guess. This, however, is sent to the USGBC for a LEED rating.

Meanwhile, BOMA continues to report that occupants are often dissatisfied with the thermal environment we provide, even with tried and true overhead delivery systems. Now we are moving to underfloor and displacement systems. It remains to be seen if they are any better at satisfying occupants. The first part of the articles describes the “rules” for occupant comfort, including ventilation and acoustics. The second part outlines issues to be considered when designing overhead (including Chilled Beams), underfloor and displacement ventilation systems. 

Failure to comply with the “rules” (some being included in buildings and others just good practice) is likely to continue to result in unhappy occupants. And, of course, we will eventually find out if any of these designs are really more efficient than others.

Authored by: Dan Int-Hout, Chief Engineer Krueger     

Monday, November 19, 2012

Ignorance is Bliss

I spent three days in upstate New York last week. I was presenting “The Basics of Air Distribution” to a Young Engineers at ASHRAE (YEA) in Rochester. I also made a few calls on design engineers. As I have said before, it is a bit surprising how few practicing engineers have a full understanding of the requirements of ASHRAE Standard 62.1 (2010) Ventilation Rate Procedure and recommendations regarding the ASHRAE Comfort Standard (55-2010). Standard 62.1 is referenced in the 2009 International Building Code, which has been adopted fully or in part by most local codes. While I expected some of the young engineers to be ignorant of the requirements, it is always a bit of a surprise when no one in the room knows the overhead heating rule: Maximum delta-t (difference between room and discharge), when heating from the ceiling is 15°F. Exceeding this has two consequences:

1.) 25% more outside air must be supplied to that zone (per 62.1 VRP). I would assume that when one is in heating mode, the outside air is cold.

2.) One can no longer claim compliance to ASHRAE 55-2012, Thermal Comfort, as there will certainly be more than 5.4F thermal stratification in the occupied zone. While Standard 55 is not a part of the Mechanical code like 62.1, it is often referenced. When it is more fully restated in code language (in progress) I expect it will become code in many locations.

Some engineering firms, to their credit, have internal standards limiting discharge temperatures. The engineers I speak to in these firms are glad to finally know why this seemingly arbitrary rule has been put in place.

Surprisingly, I still get push-back from engineers who claim they have been exceeding that for years and no one has complained. This is notwithstanding the fact that the Building Operators and Managers Assn. survey had shown dissatisfaction with the thermal environment continues to be the #1 reason for not renewing the lease in a high rise office.

I guess what they say is true: Ignorance is bliss.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Tuesday, November 6, 2012

Room Design

The K-Select program for grilles and diffusers contains a feature called “Room Design”. This feature was designed 20 years ago to help optimize the selection of overhead diffusers in large spaces, to get the best arrangement of diffusers in the space. In practice, it a great tool for large open plan spaces, for which it was designed. It does not, however, work very well in small spaces with just a few diffusers, or along a perimeter. The problem lies in applications where the space being supplied by the diffuser results in asymmetrical throws. The “characteristic room length” is defined as half the distance to the adjacent diffuser or the distance to the wall. Unfortunately, there just isn’t any way to use the Room Design feature if the distance to the wall is not the same as half the distance to the adjacent outlet. In most cases, to get the 150fpm throw to comply with ASHRAE Standard 62.1 requirement that it make it half way down the window, the diffuser must be much closer to the window than half the space to the next diffuser.

Small rooms (such as a classroom) also have the same problem with a wall about 30 feet from the window. One diffuser has to be close to the window, but will likely overthrow the one towards the opposite wall. Overthrow results in drafts at the midpoint between diffusers where the jets collide, and primary air then enters the occupied zone. (This is a bad thing). To meet the stringent sound limitations in classrooms, four diffusers are probably required, but the throws will often collide.

One proven classroom solution is to use a 3-way diffuser near the windows (with the non-open portion facing away from the window) and a 4-way behind it. The throw from the rear 4-way diffuser will simply combine with the 3-way, towards the window. The 3-way, and 4-way diffusers should be spaced so that they don’t collide in the parallel-to-the- window direction. Remember, it is usually no issue if one overthrows a wall, as the comfort zone starts a couple feet from any wall. The result will be that the diffusers won’t be located evenly, but likely somewhat more towards the side walls of the room.

The K-Select “room design” program is an interesting tool, and works well selecting diffusers for large spaces with no walls. In smaller spaces, ensure that the sum of half the diffuser spacing plus the ceiling height less 6ft is always less than the 50fpm throw values at maximum airflow.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Thursday, November 1, 2012

Acoustical Corrections

Once again, I have been asked to explain how to correct sound data for size, in this case, the length of a linear slot diffuser. The basic equation is that sound output is proportional to 10*Log (A), where A is the area of the sound generating source. In practice, this means that for doubling the area, and at the same time doubling the flow rate, the measured (and reported) sound level would be expected to increase by 10* Log(2*A), which is 3 dB. That is the rule; add 3dB for doubling the area of a sound generating source.

This comes to play when looking at the sound traveling down a duct. If the duct is split and feeds two separate spaces, any sound in the duct would be expected to be 3dB lower at it enters the space. If, of course, the two ducts exit into the same space, the sounds would recombine, thus adding the 3dB back. This effect is the only sound parameter that is independent of frequency.

For continuous slot diffusers, we provide a correction table for length, for lengths up to 10 feet. Someone wanted to know the correction for longer lengths. Applying the above formula to the base data, which is based on a 4 foot length, going to 8 feet would increase the sound by 3dB. It would go up by another 3dB at 16 feet (all assuming the flow rate per foot is kept constant). The problem is that by the time we get to 16 feet, the observer is so far away from the added length that one can no longer hear the sound being generated there. So we stop the published correction at 10 feet.

The posted sound data for continuous slots assumes that the supply plenum isn’t adding any sound. In practice, however, it is likely that there is some noise added by the supply plenum. As a rule, the larger the plenum, the less noise it will create. A “step sided” plenum (wider than the opening at the top of the diffuser) will generate less noise than one that is only as wide as the opening. A taller plenum is quieter than a shorter one, and allows for a round, rather than an oval, inlet. An oval inlet has less area than a round one with the same perimeter. As a “rule of thumb”, add 1dB for every 100 fpm velocity above 800 inlet velocity into the plenum.

Finally, insulating the inside of a plenum decreases the interior volume, raising velocities in the plenum, and negating the sound absorption of the insulation. This is one reason that plenums are insulated with thin insulation, because increasing the thickness of the insulation will likely result in more sound generation. In practice, spaces with a return air plenum are seldom faced with condensation on the plenum, as the space (and therefore plenum) dew point is almost always higher than the supply air temperature in the plenum. If the space has ducted returns, I recommend field installed external insulation on all exposed surfaces, especially in humid climates.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Wednesday, October 10, 2012

Talk to the Occupants!

Last week I spent 4 days in Southern Ontario, Canada. I spoke at three different ASHRAE chapters (London, Hamilton, and Toronto) and called on 15 design engineering firms with our local reps and the regional sales manager. In total, I saw nearly 200 folks. I asked all the same question. “What is ASHRAE’s recommended maximum delta-t when heating from the ceiling”. In spite of Standard 62’s Ventilation Rate Procedure being code in Ontario Province (they have the IMC 2009 in their code, which references the 2007 VRP), there were few who knew the answer. Sadly, several of those who knew the “rule” said they ignore it because “no one is checking”. It is a major problem if there are things in codes that no one checks.

One firm did have a design leaving air of no more than 90°F, but no one in the room knew why. They were glad to understand why this rule was in place.

Checking is surprisingly easy. Every project has a schedule that lists the design discharge temperature of all devices with heating coils. All a code official has to do is compare this data with the stated design outdoor air delivery rate. ASHRAE 62.1 (Table 6.2) clearly states that if the delivered air is more than 15°F above room temperature, they must divide the room ventilation rate by 1.2, which is a 25% increase in the outdoor air quantity, when heating. In Canada, this is likely very expensive in their cold climate. If the outdoor air damper is fixed, this increases the dehumidification and cooling demand in the summer.

In addition to Standard 62.1 compliance, when the discharge temperature is high, it is extremely unlikely that the space will meet ASHRAE Standard 55’s 5.4°F vertical temperature stratification requirement. The ASHRAE Fundamentals chapter 20 on Air Distribution states: “when the room to discharge differential exceeds 15°F, it is unlikely that the vertical temperature limitation of ASHRAE Standard 55 will be met.” When we did the overhead heating analysis in the late 70’s, we ran over 900 different perimeter conditions, and in none of the cases was there compliance to Standard 55 when the delta-t exceed 15°F.

Again, however, many in the design engineering community seem to be unaware of the fact that hot air rises and that releasing it at the ceiling will result in occupant dissatisfaction. When I bring this to the attention of many engineers, too often their response is “I’ve been discharging hot air into spaces for years and no one is complaining”. I question whether they ever actually asked anyone living in such an environment. As BOMA continues to report that the #1 reason for tenants not renewing their lease is “occupant dissatisfaction with their thermal environment,” I would conclude no one is bothering to ask the occupants.

It also appears that there is poor understanding of the need to adjust diffusers prior to balancing. One Canadian air balancer told me that they had to call the engineer to find out how to adjust the linear slots on his project, as he could find nothing in the design documentation. At least one balancer is taking the time to do it right. Sadly, he is in a minority.

It appears we have a ways to go before we manage to design comfortable spaces.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Thursday, October 4, 2012

School Acoustics

I was recently asked about the issues with meeting classroom acoustics requirements, both for the upcoming LEED V4 and many local code requirements.

The Acoustical Society Standard ANSI S12.60 is the base requirement. For most classrooms, it requires a dBA no greater than 35, which is about 26 NC, depending, of course, on the shape of the sound spectrum. This is a pretty tight requirement, especially because it doesn’t state from where the sound comes. Measurements in many schools exceed 35dBA from outside noise alone.

LEED for Schools presently has a 45 dBA limit for getting a point, again not specifying from where the sound comes.

The upcoming LEED V4, however, has a prerequisite of 40 dBA, but specifies that it is for predicted sound from the HVAC system alone; AHRI 885 is referenced as a calculation method. This is a major improvement, I believe, as one no longer needs to hire an Acoustician for this value.

Reverberation time is also specified, but in V4, it is covered by room construction details.

Now, what to do about the HVAC system? One thing is clear, if the 40 dBA requirement is to be met, there will be no mechanical equipment in the classroom and likely none above the ceiling if it contains a fan. 40 dBA is about NC 31. Putting equipment in the corridor with lined duct will be the likely solution. Fire dampers are often required at the wall penetration. There are also space issues that need to be worked out.

Displacement Ventilation diffusers are the quiet solution for air outlets, but there are challenges:

--- The air should not be any colder than 65F to avoid discomfort, but the dew point has to be much lower. The mix of ventilation air and humidity control will be a challenge. Some codes now require that the ventilation be shut off in unoccupied classrooms. This implies a separate ventilation supply, which likely has to be pressure independent VAV that allows easy demand controlled ventilation.

---  The “near zone” is likely about 4 ft at 250 cfm. This means a clear space around the diffusers, which is no problem for the ones often placed on each side of the white board, but it may be a challenge in the rear of the ever more crowded classrooms.

---  Displacement diffusers do not heat as well as other methods. Higher velocity air at the floor is likely required, or at least a baseboard or some other alternate heating solution.

---  The thermostat location in a stratified environment is problematic. Some trial and error is likely, as the ADA height requirement may not be ideal; some offset will be inevitable.

---  Finally, with many classrooms in perimeter zones, the internal heat gain in the winter will offset the heating demand. Many times, discharge delta-t’s are very low or neutral once the classroom is in use. This adds to the problem of controlling humidity without reheat.

This means that the engineer will have to look at a number of solutions. It seems that a dual duct system is a pretty good one, with ventilation air in one duct and recycled air in the other. The dual duct unit may require a heating coil in some areas, but it could be located above the classroom and still be able to meet acoustical requirements. Locate the two supply ducts one above the other, as they will be supplying air from each side.

The DOAS fan unit may be another solution, given its ability to vary the ventilation rate while separately managing heating and cooling. The unit of course, will likely have to be in the corridor.

The chilled beam has also been considered, due to its quiet nature, but openable windows offer a challenge in controlling condensation.

In any case, classroom HVAC design won’t ever be the same.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Tuesday, September 25, 2012

The Energy Code Challenge

ASHRAE Standard 90.1 2010 will be “strongly suggested” to be adopted in codes next year. (October 18 2013 is the date mandated by the US D.O.E.) ASHRAE 189 2010 will push the envelope further, and will be code in a few locations. This revision to the 90.1 Standard, as well as the new 189 Standard, will impact the way we design systems. Just recently, I attended a seminar outlining the impact of the updated standards. 52 of the 110 changes to the 2010 revision affect HVAC systems. 25K sq ft, single zone, 2 floors or less, and constant volume is a lower limit on simple systems. The Performance Rating Method, appendix G, is not mandatory, but is used for LEED comparison purposes. The Standard provides both a “Prescriptive” and a “Performance” path. This allows some flexibility in selecting the design. Air cooled equipment will be harder and harder to comply with the Standard. Water cooled equipment makes lots of alternate systems practical, including using water to move and remove heat.

Some of the issues, and how they may affect us in the air distribution wing of the HVAC industry, are as follows:

Rooftop units: Chilled Water with 5hp fans or DX > 9 tons must vary the airflow rate at low loads, which means a variable minimum outside air flow rate. Dynamic Ventilation rate control is also required.

Economizer: Previously, economizer was mandated only in milder climates, but now it has “Moved South”. Smaller rooftop units must have this capability. This will make rooftop units more complex (and expensive). It may push for VAV in many spaces that were constant volume in the past.

Reheat: Restrictions will cause engineers to rethink how they control humidity. I suspect fan powered terminal units will continue to replace single duct / reheat units as the rules tighten up.

Energy Recovery (Time for suppliers to dance a jig!): Energy recovery will become a major component of any outside air treatment system. I suspect DOAS (Dedicated Outdoor Air Systems) will get a huge boost with this.

Other requirements: Half the electrical outlets will be shut off during unoccupied periods. Roofs must reflect heat. Interior lights must be reduced. Most importantly, appendix G (which defines ‘baseline systems’) is being modified, which affects the push towards implementation of new systems.

Duct Leakage is emphasized as a major energy waster, and the new standard has very tough rules.

Finally, perimeter envelopes are highly regulated. This is especially good news as the perimeter is where most of a modern building’s energy is being consumed.

The result of all these changes will be to further reduce the energy use in the interior of buildings. This reinforces the idea that the ventilation rate requirement is the load in the interior. As perimeters get better, the ventilation rate will drive them as well. All HVAC system strategies, be they VRV, WSHP, UV, Fan Coil, DV or any other method, all require ventilation air, which is becoming the major component of the load calculations. We will come to realize that these ventilation systems need to be pressure independent as ventilation demands change within a space. In other words, VAV systems will continue to be a part of ANY HVAC system.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Wednesday, September 19, 2012

My Findings in Washington D.C

I spent the last few days in our nation’s capitol. In the schedule were a couple of speaking engagements as well as a few sales calls with our Terminal Unit Product Manager.

Interestingly, Washington D.C. has a number of very “progressive” Mechanical Engineering firms – probably more than any other city I have visited. All are very interested in “pushing the envelope” in air distribution designs and are, in one way or another, using a series fan powered VAV terminal unit as the basis of their designs, always with an ECM motor. Most are using a “Chilled Box”, or DOAS fan terminal unit, as it is sometimes called, which is a series fan box with a sensible cooling coil on the induction port. We have a white paper on this technology on the Krueger website. (

The DOAS terminal unit was the subject of the Washington D.C. chapter ASHRAE dinner talk, presented by Southland Industries. They recently finished a retrofit of the Pentagon using this technology. The talk was an excellent discussion of both the development of this strategy and the control options possible. They mentioned the beginnings of the concept and even the SSA Payment Centers with which I was involved in the ‘70’s, a precursor to this concept. My first exposure to the concept dates back to 1991. Similar to Carrier’s 36 series induction units, first installed in 1948, both the chilled beam and the DOAS fan terminal units use similar technology by combining non-condensating cooling coils and conditioned primary air to handle the sensible load in the space. That product was one of the many products I supported as terminal engineering manager back then. It is gratifying to see the excitement we observed on this technology at the many engineering firms we deal with in the D.C. area. In fact, one participant during the Q&A period stated “Since this is obviously the way we are going to be doing things in the future.”

I couldn’t agree more. The use of variable flow ECM motor technology, tied to a sensible cooling coil and a low temperature DOAS system, will likely be the most cost effective way to provide load management and ventilation to many types of spaces for some time to come.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Thursday, September 6, 2012

More on the Cost of Cool

Last week I posted some numbers from an article in the New York Times on the true cost and future of air conditioning. I have since received some interesting data on the reported energy use of some Government buildings. What I received was this:

Here are some facts to describe government building energy performance compared with the private sector. The latest CBECS Table E2A says the average energy consumption for all buildings is 91,000 Btu/SF/Yr. The 2007 Federal Government measured building energy data is:

Agency - BTU/GSF
EPA - 315,867
HHS - 313,512
DOJ - 239,558
DOE - 230,194
GSA - 74,512
Government Total - 117,495

It would appear that the GSA is doing a good job of saving energy, at least compared to the average. The same cannot be said of the other agencies. Assuming the data is accurate, the DOE, who is behind the Energy Star Programs, apparently hasn’t used their own recommendations. I attempted to break this down into average costs, but it appears it is incredibly dependent on location.

What has become apparent, however, is that interior loads continue to drop, and the main difference between different building’s energy costs is likely the envelope efficiency. Changing the thermostat setting will likely have a net zero cost on energy use, because interior loads are pretty much independent of climate, but building a better building skin is where the real savings are. Trying to squeeze energy out of interior HVAC system components is likely a fool’s errand. Energy used in the interior is converted to heat and must be removed, no matter what the setpoint, and to a great extent, no matter how it is removed. True and meaningful energy savings can be achieved by treating outside / ventilation air and avoiding over-cooling spaces when controlling humidity. Furthermore, we are realizing that the minimum ventilation rate is actually the load in interior spaces.

In summary, let’s speak to the architects about making buildings energy efficient instead of trying to squeeze the last drop of savings out of HVAC systems.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Friday, August 31, 2012

The Future of Air Conditioning

An article from the New York Times was sent to me recently. It described the problems with the energy use of air conditioning in the future as well as attempts to conserve today. Here are a few interesting facts from the article.

- A Japanese study concluded that operating an office a few degrees above “optimum” (72-75°F) resulted in a measured decrease in productivity of between 2 and 5%. If salaries run $200/SFY (square-foot-year average in developed countries), and one loses 2% productivity, that is $4/SFY. It is estimated that the average cost to heat and cool a building is about $2/SFY. It is easy to see that operating a building above a desirable setpoint is a losing proposition.

- In Japan, with the lack of nuclear power, some buildings were operated at 80°F+, to save energy. Occupants were using fans to be able to continue to work. Data showed energy use was greater than when operated at more ideal temperatures! It turns out that many buildings are mostly made up of interior spaces and the setpoint is pretty much independent of the temperature difference between the inside and outside, but wholly dependent on the internal heat gain, which must be removed when outside temperatures are above whatever setpoint is established (where outside air cannot be used to cool a building).

- Air conditioning is directly related to productivity in hot climates, where most humans now live. It was estimated that a billion (yes, a billion) folks will be entering the consumer market in the next 15 years. It is further estimated that one of their first purchases will be an air conditioner. If this proves true, providing power to run those AC units will be a real challenge; conservation and thoughts of “carbon footprints” in the US will pale in comparison to the expected increase in just over a decade.

This should be an interesting decade...

Authored by: Dan Int-Hout, Chief Engineer Krueger

Monday, August 20, 2012

Infrasonic Noise

I have been having conversations with a retired acoustical engineer in upstate New York. He has been working with folks who have been disturbed by nearby wind farm towers producing infrasonic noise.

Recent studies show that infrasonic noise is detected by the inner ear. The main unknown is our understanding of how we react to the infrasonic noise that is detected by the outer hair cells in the cochlea. It is obvious that the inner ear is much more complicated than acoustical practitioners have been taught. The old saying, "if you can't hear it, the noise can't hurt you," just isn't true.

We know that we have been hard-wired to protect ourselves when we receive certain sensory perceptions. For instance, why does one third of the population get sea-sick or experience motion sickness? What is the root cause?

If you go to Wikipedia, you will find: We take all our information from the world through our senses, and many times that comes from multiple inputs. For instance, in the case of sea sickness, the brain is processing the inputs from the following senses:

Both inner ears monitor the directions of motion in three dimensions. Our eyes observe where our body is in relation to its surroundings as well as the direction of motion. Skin pressure receptors, such as those located in the feet and seat, sense in what direction the gravitational pull affects our body, in other words: “What side is up?” Muscle and joint sensory neural receptors report which parts of the body are in motion and in which relative direction.

When feeling motion but not seeing it (for example, in a ship with no windows), the inner ear transmits to the brain that it senses motion, but the eyes tell the brain that everything is still. As a result of the discordance, the brain will come to the conclusion that one of them is hallucinating and further conclude that the hallucination is due to poison ingestion. The brain responds by inducing vomiting, to clear the supposed toxin.

In a similar manner, I don't think we understand why we react to excessive low frequency and infrasonic noise the way we do. One thing worth remembering is that in nature, when there is excessive low frequency or infrasonic noise, good things are not happening. When our forefathers and foremothers heard and felt the rumbling from thunder, tornados, and earthquakes, they knew it was time to head back to the cave. We are the survivors of those fast folks who made it back to the cave. Isn't this what happens in evolution?

As a note, most sound meters and smart phone apps can’t register infrasound.

Low frequency sound energy can cause some folks to react negatively. HVAC systems can cause low frequency sound (typically large air handlers and turbulence), leading to complaints, of which solutions are hard to come by. It is best to design systems right to avoid these kinds of problems. So to go back to the conversation with the engineer, if you have a nearby wind tower, you may understand why you or someone in your family may be bothered by its presence.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Friday, August 10, 2012

Bike Week!

I spent the last week up at the Sturgis Bike Week, where I again rode with a couple other HVAC industry guys. One was a mechanical equipment rep, the other a veteran design engineer, both from Omaha. Not surprisingly, we discussed our business. Here are some of my observations and take-aways.

1.) The economy has had a huge effect on the rally. On Monday afternoon, there were only a dozen bikes in the parking lot. Usually there are a hundred. Hewlet was as light as I have seen in 10 years and there was even available vendor space at the HD store in Rapid City. Most of the folks I met were from states surrounding South Dakota. I didn’t see a single bike going North on my way up the Thursday before, which was very unusual.

2.) We discussed specifications, which was the subject of my last blog. The rep said he sees nonsense “flawed” specs all the time. The engineer said that they don’t have time to clean up all the hundreds of pages of specifications. What ends up happening is that the rep just bids what they think the engineer actually wanted. Needless to say, a close relationship between the rep and engineer is ever so important.

3.) Ventilation air is the load. Minimums, tied to incredibly low space loads, are often too much air at 55°F. Strategies to raise the discharge in some zones needs to be considered, especially when the perimeter may well need 55°F air. The DOAS “chilled” box I discussed earlier is a great solution to this quandary. The equipment rep who sells the DOAS rooftop supply units got a bit excited and will be calling our rep to look to a partnership!

4.) Adjusting diffusers and placing them properly continues to be last on the list of concerns of most engineers and contractors. Balancing nightmares are real. Occupants who experience non-uniform environments are often very unhappy; productivity will suffer. None of the three of us could see a solution besides increasing awareness of the need for proper air distribution design. WE know how to, but most don’t.

5.) Acoustics will continue to be a big deal. I just got some info on ‘infrasound’, very low frequency noise. It has interesting effects on the inner ear. I’ll talk about that next week!

Authored by: Dan Int-Hout, Chief Engineer Krueger

Wednesday, July 25, 2012

Flawed Specifications

I was recently asked to review a specification provided by a consulting engineer for a project at a large university. I was appalled at what I found. I started to list the flaws in the specification, but after listing about a dozen, I gave up. The specification was so filled with inconsistencies and incredibly obsolete references that it made no sense to continue.

I wrote an article for the Construction Specifier magazine back in 2004 (which can be found at where I outlined a number of items found in specifications which were impossible to meet, as they were either referencing obsolete standards, required performance which could never be obtained, or were simply self-contradictory. Nevertheless, our reps and their customers are required to meet these flawed requirements.

The specification I mentioned above was worse than any I cited back in 2004. Issues found included:

- Reference to ADC 1062-R2. This standard was superseded by -R3, -R4 and finally was discontinued when the ADC stopped certifying air distribution devices in 1084. I believe -R2 was released in about 1973. It never dealt with VAV boxes in a comprehensive manner and was replaced by ASHRAE Standard 130 and ARI-880.

- There were requirements for "medium speed", an SCR speed controller, and then required ECM motors on fan boxes. There is no “medium speed” on the required ECM motors and they do not use SCR speed controllers.

- The spec called for a specific brand of pneumatic controller and then, in the same paragraph, specified DDC controllers.

- The spec then required “snap acting” actuators, “normally open”. This is a $200 add over conventional electric actuators, and “normally open” is a pneumatic control sequence.

- The acoustical portion of the spec was completely ineffective. It listed octave band sound requirements with no description of the set up which was required to measure these values, and the values themselves were unlikely to be achieved under any conditions. There was no mention of AHRI 885, the acoustical calculation specification which has been in place since 1989 and is required by manufacturer’s to use when presenting NC values in catalogs.

The challenge now is for the rep to call on the engineer and somehow point out his flawed requirements without insulting either their competence or intelligence. I have attempted to do so in the past with a large consulting engineer who continues to flatly specify a device whose published performance data is based on 1980 instrumentation and cannot be repeated under today’s Standards (ASHRAE Standard 70) and greatly underreport the device’s throws. When challenged, I was told that “fixing the specification now would, in effect, be an admission that our spec was flawed for the past 30 years”. (It was.)

At some point, consulting engineers need to review their “standard” specs for flaws, inconsistencies, and obsolete references. Meanwhile, of course, everyone will continue to bid their work, knowing that the specs  cannot be enforced.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Saturday, July 14, 2012

Slots are adjustable? Who knew!

I made a call on a design engineer firm today to do a “Basics of Air Distribution” Lunch and Learn. It turns out that I learned something (which I usually do). It seems a number of young engineers weren’t aware that linear slots were adjustable. They assumed they were placed at the window and designed to blow down. No wonder so many perimeter environments are uncomfortable and drafty.
Here’s the deal. Almost all linear slot diffusers are adjustable from horizontal to vertical. There are a few with fixed deflection. (Often, of course, these are located where they should have been adjustable.) The deflectors in most slots can be switched from deflecting left to deflecting right (and of course, down). When shipped, the factory has no idea how they should be adjusted, and often there is a wide degree of variance in the factory settings, which are also subject to shipping, installer manipulation, and general handling issues.
In practice, almost all slots should be adjusted to blow air horizontally along the ceiling. For a perimeter, the ideal location, according to research published by several manufacturers back in the late 70’s when we were actually researching this issue, is a couple feet away from the window, set to discharge some air towards the window, and some into the room. This results in a good compromise between heating and cooling performance, and has a chance of complying with the requirement in ASHRAE 62.1 that the 150 fpm throw make it to within 4.5 feet from the floor. (Failure to comply with this requirement requires a 25% increase in ventilation air to compensate for the inevitable short circuiting that will result). There is also a requirement that the discharge air not be more than 15-deg F above room temperature, or the same penalty applies. The data (and the ASHRAE Handbook) also explains that delivering air more than 15-deg F above room temperature will likely exceed the ASHRAE Standard 55’s maximum room vertical stratification limits, and void compliance with the Standard.
Failure to adjust slots in interior spaces results in cold air being directed down on occupants. This is always unacceptable. I can’t begin to tell you the number of times I have suffered in an ASHRAE technical meeting in a fancy hotel where there was a slot overhead blowing cold air directly on me. I have learned to wear a sport jacket in all ASHRAE meetings. I am tempted to take gloves and a scarf.
Adjusting a linear slot requires that the design engineer state clearly in the design documents that the slots must be adjusted, per instructions, prior to balancing. When adjusted from vertical to horizontal, the pressure drop almost always increases significantly, and must be accounted for in the system balance. While some may argue that a balancer won’t do this even if required to, it is assured that if the engineer does not state this requirement, it won’t happen.
The ball is in your court Mr. Design Engineer.
Authored by: Dan Int-Hout, Chief Engineer Krueger

Tuesday, July 3, 2012

ASHRAE San Antonio Summer Meeting

Another ASHRAE meeting is now behind us. It was in San Antonio, TX which would have been brutal in June, except that the hotel / convention center are located on the river walk, so it was actually pleasant, even in the heat of the day.

I attended a number of Technical meetings; here is a summary of issues covered and developing:

-- Standard 62.1 has been considering the issue of a minimum humidity limit. Standard 55 (Comfort) eliminated the lower limit in 2004, as there simply wasn’t enough compelling evidence to support a lower limit for comfort. Standard 62.1 is finally coming to the same conclusion on air quality data. Manufacturers of humidifiers, of course, are disappointed. There may well be data to support the need for humidification for health reasons, but the data is not very ‘robust’, and is mostly anecdotal.

-- Standard 55’s issue with a limit of 40 fpm for proper use of both the graphical and PMV methods was discussed. I accepted the assignment to provide alternate wording to the section to indicate the correct use of a 40 fpm value. It is acknowledged that any space, no matter the method of comfort control, or even passive conditions, is sure to exceed 40 fpm at some location, due to convection from local heat sources.

-- The ASHRAE Standard 130 (VAV Box Method of Test) acoustical requirement for pure tone qualification of the reverb room doesn’t make sense if the measurement requires only broadband measurements. A member will look for wording in line with current, existing standards to see if they can come up with a more logical requirement. Much of the existing VAV box data was collected in broadband qualified facilities. The wording then may be applied to Standards 70 and possibly others which have a similar issue.

-- End Reflection data was discussed again. There is an ASHRAE Journal statement from AHRI on the subject in the latest issue. We will post it our site soon. I also covered the issue in previous Blogs.

-- The ASHRAE research project 1515 has been reported to be nearly complete. The significant finding from this research conducted at Yahoo’s facility in California is how low loads are in today’s modern buildings, below 6 BTUH/SF for long portions of the day. Most interior zones are designed with about 22 BTUH/SF as a default interior load. The measured load is close to the minimum ventilation and latent control requirement, indicating that systems that manage outside air most efficiently are likely to be the most efficient and that complicated/expensive systems controlling sensible loads are likely not being used in most interior spaces.

-- I have rejoined the Technical Activities Committee (TAC), this time responsible for managing the new Multi-disciplinary Task Groups (MTG’s). These are composed of representatives from different TC’s who are coordinating cross-committee issues, meeting only by conference calls, much like the USGBC committee I have been involved with for the past 18 months. There is some very interesting work taking place here. I reported earlier that when the Handbook Committee of TC 5.3 (Air Distribution) met with the Standard 55 Committee (Comfort) in Chicago, members had to be introduced, as they hadn’t met before. I suspect there are a lot of other opportunities for interplay between committees.

The next meeting is here in Dallas, next January.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Thursday, June 7, 2012

LEED 2012 Delayed

We received a notice from the head of the USGBC that the release of the next edition of the LEED Rating System has been delayed. It was decided to give folks more time to respond to the next edition of the LEED rating system, to fully reply to those concerns and to better prepare the resource documentation for the many points that can be gained. The next edition of the LEED rating system will now be referred to as V4.0, rather than LEED 2012, and will be delayed until mid-year 2013.

This delay reflects a number of issues. LEED, like ASHRAE, is driven primarily by volunteers who also have to manage a professional career. In addition, there are several issues regarding new credits that apparently are still unresolved. While there are some issues that are still open for discussion, the ones affecting air distribution are not. I feel that the most important ones affecting our industry are as follows.

1.) ASHRAE 62.1 Ventilation Rate Procedure is a pre-requisite. There are a couple ‘zingers’ in here that are often overlooked:

-- Minimum Ventilation Rates in table 6.1 are adjusted by the air change effectiveness of table 6.2. For overhead cooling, this value is 1, but for heating from overhead, if the discharge is more than 15F above the thermostat set point, or the throw from the ceiling air outlet doesn’t come half way down the window, the minimum ventilation rate is increased by 25%. Displacement and short throw Underfloor Air Distribution (UFAD) outlets get a 25% reduction.

-- Charcoal Filters with 30% effective removal for ozone are required on outside air outlets in “EPA non-attainment zones”. At present, that is Los Angeles and Houston, but the EPA is petitioning to reduce the allowable concentration, which would greatly increase the covered areas.

-- Outside air quantities need to be measured. This will have a great effect on operation of VAV systems, which with the reduced interior loads we are seeing, will likely result in many VAV systems operating at or near 100% outside air. The good news is that in this condition, in the summer, most of the heat of the lights is discarded. The bad news is that many systems can’t actually handle the latent load at design ventilation rates. Measuring ventilation rates at each zone has the potential for significant reduction in required total ventilation rates.

2.) HVAC system acoustics has been defined as a separate item (as opposed to general room noise levels, which could include many uncontrolled noise sources). HVAC noise generation will be addressed through either the ASHRAE handbook (which is not sufficiently complete to do all the necessary sound path calculations) or more importantly, through AHRI 885, which has a spreadsheet to perform necessary calculations. This means the design engineer doesn’t have to hire an acoustical consultant for an “estimate” of the sound generation. LEED for Schools will have a 40dBA maximum requirement as a pre-requisite. Other types of spaces will have a credit available based on recommended maximum sound levels, often expressed in NC ratings.

3.) Thermal Comfort, as defined in Standard 55, will require a form be filled out which includes design clothing levels and metabolic rates of occupants. Discharging ceiling supplied, heated air that is more than 15F above room setpoint will negate the comfort credit. Room air speeds will need to be addressed, as well as vertical temperature stratification within the “occupied zone”. The use of ADPI can assist meeting this requirement for overhead air systems. Meeting the requirement for UFAD and Displacement Ventilation (DV) systems is problematic and will likely result in ‘creative’ inputs from designers that want to achieve this credit.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Tuesday, May 22, 2012

Applied Bio-physics

I spent all last week in southern Ontario, Canada. I reported earlier that I had visited Ottawa and Montreal. This trip was to several cities in S Ontario, including Toronto. I visited 11 engineering offices and met with about 80 mechanical engineers. At each location, I asked the overhead heat question: What is ASHRAE’s recommended discharge to room differential when heating from the ceiling?” Only 7 engineers knew the answer. (15 deg F). I was not surprised. The building code in Ottawa references ASHRAE 62.1 2004. So that means that 90% of the engineers would likely violate the building code when designing a system. Looking at building schedules across the US and Canada, I’d say this was about average.

When ordering VAV boxes with heat, our reps typically have to use the box schedule to select either the number of rows, or the KW. I’d estimate that 90% of the schedules have leaving air greater than 90 deg F. I would also guess that the engineer hasn’t upped the ventilation requirement by 25% as required by ASHRAE 62.1, which is often code.

Further, a discharge greater than 15 deg F above the room means the space will no longer comply with the vertical temperature stratification limits specified in ASHRAE Standard 55 (comfort). While typically not code, one cannot get the LEED comfort point with this design.

BOMA, for the past several years, reports that the #1 reason for not renewing the lease in a high rise building is occupant dissatisfaction with the thermal environment. Looking at the typical VAV box schedule, I am not surprised.

I wrote an article for the ASHRAE Journal a few years ago which is on our web site: ( All the information in that 2006 article was known in 1979. The actual business we are all involved with is “applied bio-physics”. We are providing life support systems for occupied buildings, and from my experiences, we need to be doing a better job.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Wednesday, April 25, 2012

Reviewing ASHRAE DL Speakers

I made a couple of visits to speak at Canadian ASHRAE chapters last week in Montreal and Ottawa. Everyone was extremely friendly; we had a huge attendance at both chapters - over 125 in Montreal. My talk went well, even though I am not conversant in French. It seems English is pretty well understood there.

I heard repeatedly that I was one of the best speakers they had ever had. I often get that comment, and at the same time, I hear how dull and uninteresting most ASHRAE talks are, especially with the ASHRAE Distinguished Lecture (DL) speakers. There are lots of ASHRAE DL speakers. A good number of them, in an effort to establish international credibility and recognition (ASHRAE Officially dropped the “American” name when it became simply “ASHRAE”), are not fluent in English. More, it seems, are just plain dull.

Based on all the feedback I get, the reason folks like my talks is that they are interesting, sometimes humorous, informative, and always clearly making a point. Sadly, is seems many are informative, but miss the other attributes. I understand that we are mostly engineers. Engineers have a reputation for being linear thinkers and somewhat boring - but we don’t have to be. Dull speakers result in folks not ‘bothering’ to attend meetings, which is sad, because it reduces the potential for ASHRAE Engineers to gain new learning experiences. We have to make the talks interesting if we want our message to get out. This is assuming, of course, that we have a message. Sometimes I suspect, there is no message, just a data dump, which usually results in no message being retained.

So what’s my message (you suddenly ask)? Fill out the feedback forms when an ASHRAE DL speaks. Don’t be shy about your comments. If the speaker was boring, please let ASHRAE know. If he was terrific, let ASHRAE know that too. Maybe we can weed out the bad talks so more will attend the ASHRAE sessions.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Thursday, April 12, 2012

Hospital Operating Air Systems and ASHRAE 170

There have been a number of recent calls regarding the application of the Krueger Sterilflo air curtain system and the requirements of ASHRAE 170, the Hospital Ventilation Standard. When we noted to the committee developing 170, the chair insisted that 170 did not prohibit an air curtain system, but does not endorse them either. He mentioned that there was (at that time) no ASHRAE peer reviewed paper on air curtain systems. Gerry Cook and I immediately prepared one, and it was presented at an ASHRAE seminar. The paper has been turned into a Krueger White paper which can be found at Nonetheless, the standard was published without mention of the air curtain type of hospital operating room air delivery system.

In short, the ASHRAE Standard wants to see a series of “non aspirating” laminar panels over the patient with the rest of the required airflow delivered vertically, but not necessarily from a non-aspirating diffuser. There are two realities that one must resolve in attempting to meet this standard’s requirement. The first is that no laminar panel system is truly “non aspirating”. The committee actually witnessed this fact in a manufacturer’s lab during one of the ASHRAE meetings. Laminar panels induce some air at the perimeter of their area. The second is that this induction adds to the mass of air traveling down towards the patient’s table, increasing the table velocity. There is also the “coke bottle” effect in which the colder air tends to contract the downward air pattern, also increasing the velocity at the table. Finally, the greater the number of adjacent panels, the greater the “mass effect” which also increases the local air speed at the table top. There is a great likelihood that an all panel system will exceed the velocities specified at the operating table with the required face velocities at the panels, due to all the effects listed.

An air curtain system keeps the size of the panels as small as possible (avoiding the mass effect), and the air curtain prevents the induced air from coming outside of the sterile area surrounding the patient. Tests have been conducted measuring viable particles during actual operations (a while back, as it is unlikely one could do so today with the potential liabilities that exist these days). These tests confirmed that the air curtain system is much less likely to create a non-sterile condition than a standard laminar panel system.

Meeting ASHRAE 170 is simply a matter of managing the areas of panels and air curtains to comply with ASHRAE 170.
Authored by: Dan Int-Hout, Chief Engineer Krueger

Tuesday, April 3, 2012

The New Facility is Operational!

After several months of construction, and much longer planning, the new Krueger Lab, Training, and Demonstration Center is operational. We held an Open House a couple weeks ago and the first of our revamped Krueger Institute of Technology training sessions last week. During training, we demonstrated diffuser performance with smoke at several different discharge temperatures, including tests against our unique double sided Cold Wall.

The Reverb Room is operational, nearly twice the size of the one in Tucson AZ, and with far greater airflow and sound isolation capabilities than ever before. The Throw Bay is twice as long as our previous one. In Tucson, with the highly accurate omni-directional anemometers and computer acquisition system, we discovered that one could see the effect of an opposite wall when within 20 feet of it, severely limiting the ability to get accurate throw data at high airflow rates. The ceiling is up to 12 feet high against the Cold Wall on one side, against a raised floor on the other.

We have a full scale Hospital Operating Room set up with our Sterilflo particulate control system (air curtain and downblow panel). We give thanks to Steris for their lighting and operating table donations, which help demonstrate real world velocities at the patient location to assure compliance to ASHRAE 170.

We have a “standard” 2400 cu.ft. Acoustical Demonstration Room (‘In-Situ”) for demonstrating the radiated sound transfer from ceiling plenum located sound sources, such as fan terminals and fan coils. This space has a 3ft deep plenum with open sides, as required by many acoustical demonstration specifications; in addition, we have installed rails and a winch so we can rapidly change out units as requested.

We have an operational Fume Hood Laboratory with several operational variable volume vertical sash fume hoods and Krueger ceiling displacement TAD and RadiaFlo diffusers to demonstrate face velocities at the hoods with full airflow.

We have installed a raised floor against one wall of the double-sided Cold Wall. This raised floor has several different perimeter systems as well as an array of passive interior floor high induction swirl air outlets. The air system is configured to demonstrate displacement ventilation air outlets as well. This room also has several operational chilled beam units in the ceiling and includes a humidifier to show the response of the sensible cooling coils to room humidity levels.

Finally, we have a theater type training room with seating for 40 and power and internet connections for all. This space is conditioned using displacement ventilation.

The entire space surrounding these individual areas, including the workshop, has exposed duct work, which showcase the complex system of space conditioning and laboratory air delivery. All systems are controlled through an internally programmed LabView system with over 50 pressure transducers, several theatrical smoke generators, and connected flat panel monitors with iPad accessible on-screen displays.

We look forward to having many visitors come through to see this new series of spaces for both training and product evaluation, development, and demonstration.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Thursday, March 1, 2012

End Reflection Hits The Street

The reality of the new end reflection change in discharge sound is starting to get out, so there are lots of questions. I explained the issue a few weeks back in my "End Reflection Blog". The impact is not as great as one might expect, as it only affects low frequencies, but at the same time, smaller units with small outlet ducts where sound levels are typically very low will see a greater difference. This will likely not affect terminal unit selections for most applications. The biggest issue is meeting scheduled sound levels created before the change in reported discharge sound levels.

AHRI has been tasked with releasing a white paper and putting a notice of the change in major trade magazines. That hasn’t happened yet, but it is in progress. In the meantime, most manufacturers have updated their AHRI ratings. Their on-line catalogs are slowly being updated. Most should include a reference to the new standard on the page where data is shown. I expect it will be a while before all on-line catalogs are up-to-date with the new standard. Software updates will likely take a while longer. In addition to the changed data, there are a couple of other changes required, which will likely require everyone to recompile their programs.

For a while at least, there will be two sets of discharge sound data floating around. Everyone needs to be careful which version of the data is being quoted and engineers need to begin modifying their acoustical requirements. We strongly recommend that engineers specify maximum allowed sound power using the AHRI 885 algorithms, starting with desired room sound pressure by octave band, and working backward to calculate the maximum allowed power levels. Krueger has such a spreadsheet, located on the Krueger website at

Authored by: Dan Int-Hout, Chief Engineer Krueger

Wednesday, February 8, 2012


In my last two submissions, I discussed an issue that came up at the ASHRAE meeting, which is now a couple weeks behind us. That, of course, was not all that went on there.

Here is a recap of my observations on technical matters in Chicago:

Friday: SSPC 62.1 / IAQ (I am a Consultant): The committee continues to tweak the Ventilation Standard through interpretations and addenda.

Saturday: SSPC 55 (Comfort) I am a voting Member: I reported on some of this earlier. It was interesting that when the folks from the Air Distribution Handbook Committee (TC 5.3) presented their plan (mine actually) for compliance, they all had to be introduced to the 55 members, as they had never met!

Sunday: SSPC 55, TC 5.3 Handbook/ SPC 129 (MOT ACE): The comfort folks met again in the morning, and discussed more on the methods for compliance to the Standard, which are sorely needed. The 129 committee (which is the standard for measuring air change effectiveness, the basis of table 6.2 on Standard 62.1) has determined that SF6 tracer gas is still the best choice. We will update references and present the Standard as a essentially reaffirmed, in San Antonio (next June).

Monday: Sales Meeting / TC 2.6 (Sound): We had a national sales meeting on Monday morning. It was nice to see all the reps I missed seeing last year, and of course, those I did. I reported that I saw 670 engineers last year, representing 80 Engineering firms, did 6 ASHRAE Chapter dinners, and gave 4 half day seminars on air distribution.

Tuesday: RP 1515 PMC / TC 5.3 (Member): As I reported last week, RP 1515 is looking at the Yahoo campus in California and reports very low interior loads. VAV boxes set at 30% of design (0,3 cfm/sf) were going into reheat in the afternoon. TC 5.3 is sponsoring research to look into getting ADPI data at low flows, in heating, and for other systems.

Wednesday: TRG 7 (Underfloor Design Manual rewrite): The final draft is loaded on a proprietary website and committee members are doing final edits. These should be complete in a couple of weeks. We will then hand it over to ASHRAE Staff for editorial cleanup.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Tuesday, January 31, 2012

What a Load... Room Load, That Is!

I was in Chicago for a week at the ASHRAE winter meeting. One of my tasks, as a member of a Project Monitoring Committee, is to participate in an ASHRAE Research Project (RP). ASHRAE RP 1515 is a comprehensive study of energy and occupant satisfaction at the Yahoo campus in California. The study is being conducted as a joint exercise between Berkeley, Taylor Engineering, and a manufacturer. It should be completed before the San Antonio meeting this summer.

The data so far is very interesting. In both summer and winter (this is California, so it is a mild climate), interior loads are very low. Interior loads are pretty much independent of outside conditions. The system has VAV boxes and plaque diffusers. The assumption (going in) was that at these low loads there was a potential for poor diffuser performance. What was found was the opposite. The plaque diffuser operates well at low flows and there were (little if any) complaints of “dumping” (where cold supply air falls into the space). The complaints came at full flow when the diffusers collided; it was a bit drafty at the mid-point between diffusers.

The real surprise was how low the load in the space was. The design was the “traditional” 1 CFM/SF in the interior, with the boxes set for a 30% minimum, or 0.3 CFM/SF. It was discovered that the boxes were going into reheat at that low flow! They were reset to 10% minimum with a heating setpoint at 50% (per the latest 90.1 addendum on reheat). Analysis after the reset shows that they were averaging about 0.26 CFM/SF in these interior spaces. I assume that, being Yahoo, these spaces had at least a full complement of computers in the work stations, and likely dual monitors. Yet the loads were very low.

This confirms what I have been seeing for some time. The DuPage City courthouse, that I was an observer for back in 1991, had an interior load of 0.4 CFM/SF and had issues with diffuser performance at these low loads. Nonetheless, we are seeing designs based on much higher loads. 0.5 CFM/SF is about 11 BTUH/SF with 55ºF supply air. The Yahoo buildings are likely close to 4 BTUH/SF in the interior spaces. The load is also pretty well matched to the minimum ventilation rate.

It is apparent that we need to reevaluate the way we are calculating loads and operating buildings with these low loads. Humidity control, minimum ventilation rates, and minimum effective control for VAV boxes are all tied together. The same is true for other systems, especially chilled beams. At low loads, an oversized chilled beam (which reduces first costs) is just a very expensive diffuser. The same is true for VRV, fan coil and displacement ventilation systems. Failure to understand what is happening in today’s offices can result in wasted energy as all these systems start to go into reheat to maintain acceptable space temperatures. What a load!

Authored by: Dan Int-Hout, Chief Engineer Krueger

Thursday, January 26, 2012

40 FPM, Round 2

As I mentioned in the previous blog, there is an issue with a limit of 40 fpm in the occupied zone, a limit which is likely to always be exceeded somewhere. I proposed the following to the TC 5.3 (Space Air Distribution) handbook committee last week at ASHRAE. It was determined that it belongs in the Applications Handbook, which is a couple years away before needing revision.

We presented it to the Standard 55 committee, which determined it to be inappropriate to include in the Standard, but would be useful in the upcoming Design Guide (as yet unfunded). So, I’m posting it here for comments. I have sent it to the LEED IE TAC for inclusion in the LEED 2012 reference guide, now being put together. (Send comments to

Proposed prescriptive compliance path for overhead, well-mixed air distribution, at the design stage.

6.2.1 At a minimum, for compliance to ASHRAE Standard 55 requirements in Section 6.2, when using an overhead (well-mixed) air distribution system in an office environment, the following shall be considered a set of minimum design considerations: An interior space with multiple diffusers shall be selected so that in cooling mode, the calculated Air Diffusion Performance Index, based on table 6, ASHRAE Fundamentals, Chapter 20, is at least 80% at all expected loads and flow rates. Achieving this requirement allows the designer to assume average room air speeds are less than 40 fpm, the room environment is thermally uniform, and the graphical or PMV methods of determining room setpoints may be allowed. It also assures compliance to the maximum room air stratification limit of 5.4F in the occupied zone. In addition to the requirements of, an interior space in heating mode shall be supplied by a room to discharge temperature differential of no greater than 15F. This assures compliance to the maximum room air stratification limit of 5.4F in the occupied zone. A perimeter space in heating mode shall be supplied with diffusers that result in the 150FPM throw making it to within 4.5ft of the floor along the perimeter wall, and having a room to discharge temperature differential of no greater than 15F. This assures compliance to the maximum room air stratification limit of 5.4F in the occupied zone. A perimeter space in cooling mode with both perimeter and interior diffusers, such as an open office space, shall have a calculated ADPI for the collision zone between perimeter and interior diffusers of at least 80%, assuring no objectional drafts will be experienced by the occupants. All diffusers shall be located so as to avoid obstructions that would cause the discharge jet from the diffuser being directed into the occupied space. This assures that no excessive air speeds will be encountered by the occupants, which would void the use of the PMV or Graphical methods to determine a proper room setpoint.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Friday, January 13, 2012

The 40 FPM Quandry

We, as well as some of our competitors, have had engineers call and ask if we can guarantee that our air distribution system can guarantee to achieve less than 40 fpm at all points in the occupied zone, in order to meet LEED thermal comfort requirements. The fact is that no overhead system can guarantee meeting this requirement. The requirement comes from a statement in ASHRAE Standard 55 that the Graphical Method, "Figure specifies the comfort zone for environments that meet the above criteria and where the air speeds are not greater than 0.20 m/s (40 ft/min)." The graphical method is the most used method in the standard, as it is the easiest.

In another section of the 55 Standard; however, it states "the designer shall decide the proper averaging for air speed for use in the Graphical Method (". One method (in fact the only method) of estimating room air speeds is to use the ADPI methodology described in the ASHRAE Fundamentals Handbook, chapter 20. If the ADPI is calculated to be at least 80%, the average room air speed can be assured to be less than 40fpm. In fact, it is usually no higher than 30 fpm. This relationship will be included in an addition to the Handbook in the near future.

In addition, an ADPI >80% will also assure that there is less than the maximum 5.4F vertical temperature stratification in the occupied zone, another Standard 55 requirement. Thus, if the engineer chooses to use ADPI as the method of estimating air speed, he will be allowed to use the graphical method of determining design temperatures and assuring compliance to the stratification requirement as well.

Krueger has the best, and easiest, ADPI selection methods in the industry, including ADPI selection graphs in the catalog and an easy to use ADPI calculator and printable graph output in K-Select. We also provide the only available Standard 55 Graphical Method computer program as a free download on our website. (

Authored by: Dan Int-Hout, Chief Engineer Krueger

Monday, January 9, 2012

Top 10 Predictions for 2012

1. LEED 2012 will be released in the fall, and it will result in a number of changes in the way building HVAC systems are designed, both for energy savings and acoustics.

2. The ASA Classroom acoustics recommendation (essentially NC=26, ANSI 12.60) will become a default requirement for new classroom design. The LEED prerequisite will require that either the ASHRAE handbook, or AHRI 885 will be designated as a primary method of estimating sound levels in schools and other places as well; engineers will discover that the 885 spreadsheet is a much easier approach.

3. ASHRAE Thermal Comfort Standard (55-2010) will include mandatory language only for the calculations employed. A set of default compliance paths will be developed and put out for public review.

4. The updated ASHRAE UFAD Design Guide will come out of committee at the summer meeting.

5. Displacement Ventilation will continue to see significant use in classrooms, due to its low sound generation.

6. Chilled Beams will continue to be used by innovative engineers. While I hope to see some validation of the energy consumption savings, I predict we won’t.

7. BOMA will continue to state that the number 1 reason for not renewing the lease in high rise buildings is “occupant dissatisfaction with the building environment” (ie: comfort). I still see new designs with designed discharge temperatures for overhead systems in excess of 100°F, which leads to significant stratification.

8. The market for HVAC components will again be relatively flat with local ups and downs.

9. VAV overhead air distribution will continue to be (by far) the most used system in new buildings. ADPI will be endorsed by the ASHRAE Fundamentals as a method of predicting both thermal uniformity (complying with the vertical stratification requirement of standard 55), and that the average room air speed is less than 40fpm, allowing the graphical and PMV calculations of Standard 55.

10. Sadly, the Cowboys will continue to disappoint their fans.

Authored by: Dan Int-Hout, Chief Engineer Krueger

Tuesday, January 3, 2012

2011 Prediction Recap

It’s time for the 2011 recap. So how did I do with my January 2011 top 10 predictions?

1. LEED 2012 will be approved, pretty much as it is in its first public review.
Final public review will be out in a couple of months. I give this a 9 score.

2. The ASA Classroom acoustics recommendation (essentially NC=26, ANSI 12.60) will become a requirement in many local codes and may be a part of the next IBC. It will also be a requirement of the ADA governing bodies, forcing a redesign of many school HVAC systems.
Still on track. AHRI 885 is designated as a primary method of estimating sound levels in schools and other places as well, in LEED 2012. I give this a 9 score.

3. Over my objections, the ASHRAE Thermal Comfort Standard (55-2010) will be rewritten into code language. I predict this will result in gross misuse and misinterpretation of the requirements and will make lots of money for trial lawyers. (I hope I’m wrong on this one.)
The Standard is being revised (by addenda) to make calculations mandatory, with an informative appendix containing recommendations. I give this a 5 score.

4. With any luck, we will get a rewritten ASHRAE UFAD Design Guide out of committee at the summer meeting. The current manual has been pulled from the ASHRAE bookstore following complaints, most noticeably from the GSA, over lack of discussion of the potential negative issues with this design concept.
Final first draft has been posted and is hoped to be voted on in Chicago. I give this a score of 8.

5. Displacement Ventilation will see significant use in classrooms (it’s quiet, see item #2 above).
This trend continues. This is a 10.

6. Chilled Beams will continue to be the “darling” of innovative engineers. We will not see validation of the calculated energy savings from this relatively new technology. That will not, however, prevent estimates of significant energy savings, resulting in LEED points.
Still no peer reviewed data on energy savings with this technology. Developing AHRI and ASHRAE test methods should put to rest some unfounded performance claims. This is a 10.

7. For the umpteenth year in a row, BOMA will state that the number 1 reason for not renewing the lease in high rise buildings is “occupant dissatisfaction with the building environment” (ie: comfort). See item #3 above.
BOMA confirms that 2010 was another year in which occupants are not happy with their environment. This gets a 10.

8. The market for HVAC components will continue to be relatively flat with local ups and downs.
Pricing was a challenge in 2011, and will continue in 2012. I give this a 10.

9. VAV overhead air distribution will continue to be (by far) the most used system in new buildings.
Most designs continue with the tried and true overhead VAV designs. This is a 10.

10. The Cowboys will continue to disappoint their fans.
Fortunately, we won’t have to watch them for another 6 months. I give this a BIG 10.

Total score: 91. (out of a possible 100).

Let’s see if I can do as well in my predictions for 2012. Top 10 for 2012 next week!

Authored by: Dan Int-Hout, Chief Engineer Krueger